1. Field of the Invention
The present invention relates generally to hydrostatic bearings which can withstand a radial or axial load exerted, by means of an element rotating at high speed, such as for example, the spindle of a machine tool, upon a stationary element or vice versa, and more particularly to machine tool spindles wherein the loads are small however the increased cushioning effects afforded by means of the hydrostatic bearings serve to maintain a high degree of rigidity and thus firm guidance of the spindle which makes possible the precise machining of surfaces by means of the machine tool.
2. Description of the Prior Art
Hydrodynamic bearings having lobes have long been used in the form of journal bearings having radial projections at right angles to which the clearance between the rotating shafts and the bearings were very small. Such bearings are sometimes provided with means for adjusting the clearance by deformation in order to compensate for wear. It is also well known however that such bearings having lobes exhibit the drawback or disadvantage of high energy dissipation when the shaft speed increases, the result being a substantial rise in the temperature of the cushioning fluid which effect has many deleterious disadvantages.
Hydrostatic bearings having cavities or pockets bounded by sealing lands approaching to within a very small distance from the rotating shaft are also known and within such bearings of conventional type each pocket is individually supplied with pressurized cushioning fluid by means of a restricted passageway. It is known that the pressure within each of the pockets of such a bearing, that is, the cushioning effect of the bearing, decreases as the clearance between the sealing lands and the peripheral surface of the rotating shaft increases, and such hydrostatic bearings of the conventional type exhibit the drawback or disadvantage that the restrictions are liable to become blocked whereupon external pumping means become necessary. In addition, it is likewise known that these conventional hydrostatic bearings are functionally limited with respect to operating speed. In fact, if one considers the ratio of the power dissipated within the bearing by the shearing of the film of oil between the bearing and the rotating shaft to the pumping power, it can be shown that in practice this ratio must never exceed a value of the order of 3 to 5.
The pumping power, as is known, is equal to the product of the hydrostatic flow multiplied by the pump pressure. It has been found that as the speed of rotation of the shaft or spindle cushioned by the hydrostatic bearing increases, there is a large increase in the power dissipated within the bearing. It can also be shown that, at constant temperature, the power dissipated within the bearing as a result of the shearing of the oil film increases as the square of the speed of rotation of the spindle. In effect, as the pumping power is independent of the speed of rotation of the spindle, the flow of oil to the hydrostatic bearing is also independent thereof, the result being that the temperature of the exiting oil keeps increasing as the speed of rotation of the spindle increases, such being expressed as follows: EQU T = a + kV.sup.2
wherein T is the temperature of the exiting oil, V is the speed of rotation of the spindle, and k is a constant dependent upon the geometry of the bearing and the characteristics of the oil. It follows that a conventional hydrostatic bearing designed for low speed operation cannot be used for high speed operation which constitutes a very serious handicap which naturally limits the use of conventional hydrostatic bearings.
This disadvantage could theoretically be eliminated by increasing the pumping power, pressure, or the flow of fluid within the bearings. However, for reasons of reliability and convenience, an increase in pressure is limited in practice and an increase in pumping power can be obtained only by increasing the flow of fluid within the bearing, that is, by increasing the clearance between the bearing and the rotating shaft. Such large operating clearances entail higher degrees of pumping power, thus causing additional heating of the pressurized oil which can pose delicate practical problems if the volume of the oil is not sufficiently large. Finally, the supply of fluid to the bearings and the return of the fluid from the bearings present great difficulties, particularly because the size of the hydraulic circuitry required increases rapidly.
The self-pressurizing hydrostatic bearing described within French Pat. No. 2,157,107, issued May 7, 1973 offers significant improvement over conventional hydrostatic bearings, the cushioning pockets being supplied with fluid by means of channels which are bounded, as are the pockets, by means of sealing lands integral with an element of the bearing and extending to within a very small distance from the spindle surface. Compared to hydrostatic bearings of conventional design, the self-pressurizing bearing permits a reduction in the temperature rise with speed. More particularly, it can be shown that the pressure within the cushioning pockets is proportional to the speed of rotation of the spindle. if it is accepted that in practice the viscosity of the fluid remains constant, it follows that the temperature of the return fluid is an increasing linear function of the spindle speed, and such may be expressed as follows: EQU T = a' + K'V
wherein T is the temperature of the exiting fluid, V is the rotational speed of the spindle, and K' is a constant dependent solely upon the geometry of the bearing and the characteristics of the oil. This then represents a considerable advance beyond conventional hydrostatic bearings.
Still, with the temperature nevertheless increasing as a function of speed, limits will again be reached very rapidly. If it is desired for example to increase the speed of rotation of the spindle, the preceding problems will again arise and it will be necessary to widen the clearance between the spindle and the bearing. Since this type of bearing is self-pressurizing, the clearance entails a low-speed limit upon its use because the rigidity of the bearing, that is, the quotient of the change in pressure within the bearing divided by the change in operating clearance, is proportional to the speed of the spindle and is thus liable to become insufficient when the speed falls below a predetermined lower limit. Such self-pressurizing hydrostatic bearings are therefore not universally applicable and cannot be recommended for spindles rotating at a substantially high speed of rotation.
Furthermore, the supply of fluid to the cushioning pockets is, in accordance with such prior art invention, by way of a channel adjoining each pocket. Such a configuration does not allow the supply channel to have a substantial length which limits the pressure within the pockets since such is directly proportional to the length of the channels as is demonstrated within the description of the aforenoted prior art invention.